Hydraulic continuously variable transmission for use on vehicle

ABSTRACT

A hydraulic continuously variable transmission for use on a vehicle comprises a hydraulic variable pump, which is driven by an engine, and a hydraulic variable motor, which is driven by oil discharged from the hydraulic variable pump to drive wheels. In addition, this transmission includes shift position detecting means, which detects a shift position, and vehicle stop detecting means, which detects whether the vehicle is in halt. In this transmission, when the vehicle stop detecting means detects that the vehicle is in halt and the shift position detecting means detects that the shift position is in a forward or rearward drive position, the discharge capacity of the hydraulic variable motor is set to a maximum, and the discharge capacity of the hydraulic variable pump is controlled to a predetermined capacity (first predetermined capacity) in correspondence with the shift position. Thereby, the vehicle is prevented from relapse phenomenon, i.e., being dragged in the direction opposite to the direction set with the shift lever when the vehicle is stopped on a slope.

RELATED APPLICATIONS

This application claims the priority of Japanese Patent Application No. 9-137872 filed on May 12, 1997, which is incorporated herein by reference.

FIELD OF THE INVENTION

The present invention relates to a hydraulic continuously variable transmission, which can be used on a vehicle. This hydraulic transmission comprises a hydraulic variable pump and a hydraulic variable motor, and the pump being driven by the engine of the vehicle delivers the hydraulic oil to actuate the motor, which will drive the wheels.

BACKGROUND OF THE INVENTION

Such hydraulic transmissions have been well-known and put into use in automobiles and other vehicles. For example, Japanese Laid-Open Patent Publication No. H6(1994)-42635 discloses an automotive mowing tractor which comprises a hydraulic swash plate pump for variable delivery and a hydraulic motor with fixed displacement. In this tractor, the pump, which is driven by the engine, delivers the hydraulic pressure to the motor, which rotates the wheels. Furthermore, this tractor is provided with a seesaw pedal, which is stepped and operated by the driver with a foot. When this pedal is stepped forward, the swash plate is swiveled to a “forward drive direction”, and the vehicle is driven forward. Likewise, when this pedal is stepped rearward, the swash plate is swiveled to a “rearward drive direction”, and the vehicle is driven backward.

This pedal is designed to automatically return to a neutral position when it is not stepped, i.e., when it is released. With the pedal at the neutral position, the angle of skew of the swash plate is zero, and the pump discharges no oil. In this condition, the rotation of the motor is locked because the pump does not allow any oil to flow. This locking is so complete that any attempt to rotate the motor, which must force the oil to flow through the pump, is futile. Thus, without operating the brake, the driver can prevent the wheels from rotating by maintaining the skew angle of the swash plate of the pump at zero and locking the rotation of the motor.

If the vehicle is on a slope, for example, and the swash plate of the pump is operated to make the skew angle of the pump to zero in order to lock the rotation of the motor, then the motor is subjected to a torque which is transmitted through the wheels of the vehicle because of the component force resulting from the weight of the vehicle that acts to move the vehicle in the downward direction parallel to the slope. In this condition, as long as there is no leak in the hydraulic circuit, which includes the pump and the motor, the rotation of the motor is kept being locked. However, if any leak occurs, then the motor will rotate slowly at the speed which corresponds to the rate of the oil leak. In this leaky condition, even if the driver tries to stop the vehicle by setting the skew angle of the swash plate to zero, the vehicle moves slowly down the slope.

This problem is not very serious if the direction of unintended slow move of the vehicle is forward, and such creeping is often observed in automobiles which are equipped with an automatic gear shift. However, if the direction of this unintended move is rearward, then the problem is serious. For example, if the vehicle moves back slowly when the driver stops the vehicle on an up-hill slope with the shift lever set at the D range (a forward drive), then it is a dangerous condition.

SUMMARY OF THE INVENTION

The present invention is to solve the above mentioned problems. It is an object of the present invention to make sure of the operation of hydraulic lock in the standstill of a vehicle on a slope, the vehicle being equipped with a hydraulic continuously variable transmission.

It is another object of the present invention to prevent a vehicle which is stopped on a slope in hydraulic lock from moving downward slowly against the direction set with the shift lever (prevention of relapse phenomenon).

It is yet another object of the present invention to make sure the standstill of a vehicle in hydraulic lock and at the same time to improve the fuel efficiency of the engine of the vehicle.

It is still another object of the present invention to improve the fuel efficiency of the engine of a vehicle which is stationary in braking operation and to realize the vehicle's smooth start when the brake pedal is released.

In order to achieve these objects, the present invention embodies a hydraulic continuously variable transmission comprising a hydraulic variable pump, which is driven by the engine, and a hydraulic variable motor, which is driven by oil discharged from the hydraulic variable pump to drive the drive wheels. This transmission further comprises shift position detecting means, which detects a shift position that is set with the shift lever operated by the driver; vehicle stop detecting means, which detects whether the vehicle is in halt or not; pump capacity controlling means, which controls discharge capacity of the hydraulic variable pump; and motor capacity controlling means, which controls discharge capacity of the hydraulic variable motor. In this transmission, when the vehicle stop detecting means detects that the vehicle is in halt and the shift position detecting means detects that the shift position is in a drive position, the pump capacity controlling means controls the discharge capacity of the hydraulic variable pump to a first pump capacity in correspondence with the shift position, which is detected by the shift position detecting means, and the motor capacity controlling means controls the discharge capacity of the hydraulic variable motor to a predetermined motor capacity (>0).

With this construction, when the vehicle is stopped with the shift position set for a forward or rearward drive, the capacity of the hydraulic motor is set at a predetermined capacity (e.g., the maximum capacity), and the capacity of the hydraulic pump is set at the first pump capacity. As a result, for example, a little amount of hydraulic oil which corresponds with the set capacity is supplied to the motor that is idling while the vehicle is in halt. Therefore, this amount of oil, which is supplied to the motor, is set to such amount as to supplement the amount of oil leaking in the transmission when the vehicle is stopped on an allowable slope (without any braking operation) with the hydraulic motor receiving a driving torque. In this way, the hydraulic motor is locked, and the vehicle is retained stationary on the slope.

In this case, a problem is that the vehicle may move in the direction opposite to the direction set with the shift position. For example, the vehicle may move backward even though the shift position is set for a forward drive. However, the capacity of the pump is controlled to prevent such relapse movement.

Furthermore, if the vehicle stop detecting means detects that the vehicle is in halt, the shift position detecting means detects that the shift position is in a forward or rearward drive position, and the brake operation detecting means detects the operation of the brake; then it is preferable that the pump capacity controlling means control the discharge capacity of the hydraulic variable pump to a second pump capacity, which is smaller than the first pump capacity. When the brake of the vehicle is operated, there is no relapse movement, so the skew angle of the pump swash plate can be set to zero. However, the skew angle of the swash plate is set to create the second pump capacity, which is smaller than the first pump capacity, so that the vehicle can start smoothly from the halt when the brake is released. Thereby, the engine drive is decreased while the brake is being operated, thus improving the fuel efficiency as well as realizing the vehicles smooth switching to the start-up control.

Further scope of applicability of the present invention will become apparent from the detailed description given hereinafter. However, it should be understood that the detailed description and specific examples, while indicating preferred embodiments of the invention, are given by way of illustration only, since various changes and modifications within the spirit and scope of the invention will become apparent to those skilled in the art from this detailed description.

BRIEF DESCRIPTION OF THE DRAWINGS

The present invention will become more fully understood from the detailed description given herein below and the accompanying drawings which are given by way of illustration only and thus are not limitative of the present invention and wherein:

FIG. 1 is a schematic view showing the construction of a hydraulic continuously variable transmission, which is used on a vehicle, according to the present invention;

FIG. 2 is a graph showing relations between the skew angles of the swash plates of the pump and the motor of the hydraulic transmission and the total speed ratio of the transmission;

FIG. 3 is a hydraulic circuit diagram showing the closed hydraulic circuit of the transmission and a circuit for providing a control pressure;

FIG. 4 is a schematic, sectional view of a regulator valve, which is used in the above mentioned control pressure circuit;

FIG. 5 is a schematic, sectional view of a modulator valve, which is used in the control pressure circuit;

FIG. 6 is a graph showing a relation between the control current (I) and the control pressure (Pc) of a first linear solenoid valve;

FIG. 7 is a graph showing a relation between the control pressure (P) and the line pressure (PL) of the regulator valve;

FIG. 8 a schematic, sectional view of a shuttle valve, which is used in the control pressure circuit;

FIG. 9 is a schematic, sectional view of a high-pressure release valve, which is used in the control pressure circuit;

FIG. 10 is a graph showing relations between the high relief pressure (PH) of the high-pressure release valve and the control current (I) of second and third linear solenoid valves;

FIG. 11 is a schematic view showing the surface configuration of the valve plate of the hydraulic pump and the location of variable notch valves;

FIG. 12 is a schematic, sectional view of the variable notch valve;

FIG. 13 is a hydraulic circuit diagram showing a circuit which controls the skewing angles of the pump and the motor;

FIG. 14 is a schematic diagram of a control device which controls the operation of the continuously variable transmission of the present invention;

FIG. 15 is a main flow chart showing the control flow of the above control device;

FIG. 16 is a graph showing relations between the vehicle speed and the engine throttle (i.e., accelerator opening) and the target rotational speed of the engine;

FIG. 17 is a subflow chart showing control determinations, which are part of the control flow shown in the main flow chart;

FIG. 18 is a subflow chart showing angular determinations, which are part of the control flow shown in the main flow chart;

FIG. 19 is another subflow chart, which shows angular determinations made in the control flow shown in the main flow chart;

FIG. 20 is a side view of a creep lever and an EXC switch, which are provided at the driver seat.

FIG. 21 is a graph showing a relation between the amount of displacement of the creep lever and the target skew angle of the swash plate of the pump.

FIG. 22 is a list showing the initial skew angles of the swash plates of the pump and the motor;

FIG. 23 is a subflow chart showing the calculation of the target skew angles of the swash plates (Step D12);

FIG. 24 is a graph showing a relation between angular deviation and the current applied to a fourth linear solenoid valve to clear target deviation;

FIG. 25 is a subflow chart, which depicts Ne control flow in the control flow shown in the main flow chart;

FIG. 26 is a graph showing relations between the skew angles of the swash plates of the pump and the motor of the hydraulic transmission and the total speed ratios, and this graph also depicts control area;

FIG. 27 is a subflow chart showing control determinations for the control area, which is part of the Ne control subflow shown in FIG. 25;

FIG. 28 is a comprised of subflow charts, FIG. 28(a) through 28(d) showing control area determinations, which is part of the subflow shown in FIG. 27;

FIG. 29 is a graph showing a relation between the speed gain coefficient of speed change which is determined in the Ne control subflow shown in FIG. 25 and the vehicle speed;

FIG. 30 is a subflow chart showing the control of the swash plate of the pump, which is part of the Ne control subflow shown in FIG. 25;

FIG. 31 is a graph showing a relation between the current inducing the target skew angle of the swash plate of the pump and the speed of speed change;

FIG. 32 is a graph showing a relation between the target current (ICMDp) applied to the fourth linear solenoid valve and the control pressure PC;

FIG. 33 is a subflow chart showing the control of the swash plate of the motor, which is part of the Ne control subflow shown in FIG. 25;

FIG. 34 a graph showing a relation between the current inducing the target skew angle of the swash plate of the motor and the speed of speed change;

FIG. 35 is a graph showing a relation between the target current (ICMDm) applied to a fifth linear solenoid valve and a control pressure MC;

FIG. 36 is a subflow chart showing a lock-up control, which is part of the Ne control subflow shown in FIG. 25;

FIG. 37 is a graph showing a relation between the current inducing the lock up and the speed of speed change;

FIG. 38 is a graph showing a relation between the target current applied for the lock up and the control pressure LC for the lock up;

FIG. 39 is a subflow chart showing a relief control process, which is part of the control flow shown in the main flow chart;

FIG. 40 is a subflow chart showing a relief control process, which is part of the control flow shown in the main flow chart;

FIG. 41 is a list showing the pressures of first and second lines which correspond to the drive direction and the acceleration or deceleration of the vehicle;

FIG. 42 is a graph showing a relation between the skew angle of the swash plate of the pump and the high relief pressure while the vehicle is in acceleration;

FIG. 43 is a graph showing a relation between the vehicle speed and the target high relief pressure while the vehicle is in deceleration;

FIG. 44 is a list showing the setting for the high relief pressure in correspondence with a specific condition;

FIG. 45 is a graph showing a relation between the target high relief pressure and the target control current applied to the second and third linear solenoid valves to achieve this pressure;

FIG. 46 is a subflow chart that shows variable notch control, which is part of the control flow shown in the main flow chart;

FIG. 47 is a list showing relations among the discharge of the pump and the high pressure generated and the open or close control of the notch valve;

FIG. 48 is a graph showing a relation between the skew angle of the swash plate and the discharge of the pump;

FIG. 49 is a skeleton diagram showing a power transmission system, which is applied to the vehicle equipped with the continuously variable transmission of the present invention;

FIG. 50 is a subflow chart showing a regulator control process, which is part of the control flow shown in the main flow chart;

FIG. 51 is a graph showing a relation between the control current of the first linear solenoid valve and the pressure of an oil passage in higher pressure; and

FIG. 52 is a graph showing a relation between the control current of the first linear solenoid valve and the line pressure PL.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Construction of the transmission

FIG. 1 shows a continuously variable transmission T. As seen from the figure, this transmission T can be used either for front wheel drive or for rear wheel drive, and it is a hydromechanical transmission, which comprises a mechanical unit 1 for drive transmission and a hydrostatic unit 2 for continuous speed variation. An engine E to drive this transmission is provided opposite the hydrostatic unit 2, making the mechanical unit 1 being positioned therebetween.

The mechanical drive transmission unit 1 comprises a power distribution mechanism 3, a power transmission mechanism 4, and a final speed reduction mechanism 5 in a first casing 1 c. The power distribution mechanism 3 comprises the input shaft 9 of the transmission, which is connected with the output shaft 7 of the engine E through a torque damper 8; a carrier 11, which is directly connected to the input shaft 9; and a pump input shaft 10, which faces the carrier 11 and extends coaxially and opposite to the input shaft 9. In addition, a plurality of pinion spindles 12 are integrally provided on the carrier 11 in such a way that the pinion spindles revolve around the input shaft 10. On each of these pinion spindles 12, a pair of large and small pinions 13 and 14 which are connected to each other in integration are pivotally provided. A small sun gear 15 with a relatively small diameter which meshes with the large pinions 13 having a relatively large diameter is provided on the input shaft 10 of the pump. In addition, a large sun gear 16 with a relatively large diameter which meshes with the small pinions 14 having a relatively small diameter is provided pivotally on the input shaft 10.

An interdrive gear 18 is provided pivotally on the pump input shaft 10 and integrally connected with the large sun gear 16, and an interdriven gear 19, which meshes with the interdrive gear 18, is provided on the output shaft 17 of the motor. These drive and driven gears 18 and 19 constitute the above mentioned power transmission mechanism 4. Furthermore, a final drive gear 20 is provided on the motor output shaft 17, and a final driven gear 21 which incorporates a differential mechanism 22 meshes with the final drive gear 20. These final drive and driven gears constitute the above mentioned final speed reduction mechanism 5. From the differential mechanism 22, a right axle 23R and a left axle 23L extend in the respective opposing directions, and through these axles 23R and 23L, the mechanical power is transmitted to drive the right and left wheels (not shown).

The hydrostatic unit 2 for continuous speed variation comprises a hydraulic plunger type swash plate pump 24 for variable delivery, a hydraulic plunger type swash plate motor 25 for variable displacement, and a control block 27, which incorporates a hydraulic closed circuit 26 interconnecting the pump 24 and the motor 25. The control block 27 is provided between the mechanical drive transmission unit 1 and the pump 24 and the motor 25, being attached to a side of the mechanical drive transmission unit 1, such that the control block 27 pivotally supports the pump input shaft 10 and the motor output shaft 17.

The hydraulic pump 24 comprises a pump cylinder block 28 having a plurality of cylinder bores 29 around the rotational axis thereof, a plurality of pump plungers 30, which are provided slidably in the cylinder bores 29, and a pump swash plate 32 with variable skew, which slidably abuts on the shoes 31 that are pivotally attached to the tops of the plunger 30. The pump cylinder block 28 is coaxially connected to the pump input shaft 10 and in rotatably contact with a valve plate face 27 a of the control block 27.

The pump swash plate 32 is swiveled about a trunnion 33, which is provided orthogonal to the pump input shaft 10 (i.e., perpendicular to the plane of FIG. 1). This swash plate can be oriented at any position between an upright position, where the swash plate is perpendicular to the pump input shaft 10 as shown in real line in the figure (the skew angle of the pump swash plate is zero NNNα=0), and predetermined right and left maximally tilted positions (α=αR(MAX) and α F(MAX)), where the swash plate is swiveled rightward at a predetermined maximum angle or leftward at the predetermined maximum angle as shown in dotted line in the figure. By swiveling this swash plate, it is possible to adjust the stroke of the pump plungers 30, which reciprocate in the cylinder bores when the pump cylinder block 28 is rotated. While the pump input shaft 10 is rotated by the engine E, if the swash plate is at the upright position, then the amount of stroke is zero, and the pump discharges no oil. While the swash plate is being swiveled toward the maximally tilted position with an increasing skew angle, the amount of stroke gradually increases, and the pump discharges in corresponding, increased amount. The direction of the oil flow is dependent on the direction of the tilt of the swash plate either toward the right maximally tilted position or toward the left maximally tilted position. As described in detail later, this tilt of the swash plate determines the direction of the vehicle whether it moves forward or rearward.

The hydraulic motor 25 comprises a motor cylinder block 34 having a plurality of cylinder bores 35 around the rotational axis thereof, a plurality of motor plungers 36, which are provided slidably in the cylinder bores 35, and a motor swash plate 38 with variable skew, which slidably abuts on the shoes 37 that are pivotally attached to the tops of the motor plunger 36. The motor cylinder block 34 is coaxially connected to the motor output shaft 17 and in rotatably contact with a valve plate face 27 a of the control block 27. As seen, the hydraulic motor 25 is a variable displacement motor with plungers and a swash plate.

The motor swash plate 38 is swiveled about a trunnion 39, which is provided orthogonal to the motor output shaft 17 (i.e., perpendicular to the plane of FIG. 1). The swash plate can be oriented at any position between an upright position, where the swash plate is perpendicular to the motor output shaft 17 as shown in dotted line in the figure (the skew angle of the motor swash plate is zero β=0), and a predetermined maximally tilted position (β=β (MAX)), where the swash plate is swiveled rightward at a predetermined maximum angle as shown in real line in the figure. By swiveling this swash plate, it is possible to adjust the rotation of the motor cylinder block 34 while a predetermined hydraulic pressure is applied by the pump 24 (with a constant flow of oil). When the swash plate is at the upright position, the motor cylinder block 34 is retained stationary. As the swash plate is swiveled toward the maximally tilted position with an increasing skew angle, the rotational speed of the motor cylinder block 34 gradually increases. The motor swash plate 38 can be swiveled only between the upright position and the clockwise tilted position as shown in the figure.

A second casing 2 c accommodates the pump 24 and the motor 25, and the second casing is combined to the control block 27, which is combined to the first casing accommodating the mechanical unit 1.

Operation of the transmission

Now, the operation of the above continuously variable transmission is described. When the engine E is started, the engine output is transmitted from the output shaft 7 of the engine through the torque damper 8 to the transmission. In this condition, the input shaft 9 and the carrier 11 of the transmission are driven at the same rotational speed as the output shaft 7 of the engine. As the carrier 11 is driven, the power from the engine is divided into two portions and transmitted through the large and small pinions 13 and 14 to the small and large sun gears 15 and 16.

With reference to FIG. 2, which shows relations between the skew angles α and β of the swash plates of the pump 24 and the motor 25 and the total speed ratio e of the transmission, the above mentioned division of the power is explained. The total speed ratio e is the ratio of the output rotational speed to the input rotational speed of the transmission T, and it is given in Equation (1). In FIG. 2, the ordinate represents the skew angles of the swash plates of the pump and the motor. The rightward tilts of the swash plates are plotted in the positive quadrants, and the leftward tilts are plotted in the negative quadrants. The abscissa represents the total speed ratio e, and the speed ratio in forward drive is plotted in the positive quadrants, and the speed ratio in rearward drive is plotted in the negative quadrants. In the figure, the skew angle of the swash plate of the pump is shown in real line, and that of the motor is shown in dotted line.

Total speed ration e=(No)/(Ni)  (1)

where Ni is the rotational speed of the input shaft 9 of the transmission, and No is the rotational speed of the final driven gear 21.

When the pump swash plate 32 is at the upright position (α=α) and the motor swash plate 38 is at the maximally tilted position (β=β β (MAX)), the pump cylinder block 28 is rotatable freely, and the pump discharges no oil. The motor cylinder block 34 is hydraulically locked and retained stationary because no oil is supplied there from the pump 24. As a result, the large sun gear 16 and the interdrive gear 18 are stationary, and the small sun gear 15 rotates freely (together with the pump input shaft 10 and the pump cylinder block 28, which are connected to the small sun gear) as the carrier 11 rotates. In this condition, the engine output is wasted in idling, and no power is transmitted to the right and left axles 23R and 23L. This condition is depicted by vertical line a in FIG. 2. In this condition, the total speed ratio is zero (e=0), and the transmission T has an infinite speed change ratio.

This condition is selected when the shift lever, which is operated by the driver at the driver seat, is set at the D or R range, which is a vehicle drive range. If the shift lever is set at the P or N range, then the skew angle of the motor swash plate is controlled to become zero (β=0). As a result, the motor cylinder block 34 also becomes freely rotatable, and a neutral condition is established.

When the pump swash plate 32 is swiveled clockwise from this condition, the discharge of oil from the pump 24 starts and increases in correspondence with the increase of the skew angle of the swash plate. The oil discharged from the pump is supplied to the motor 25 to drive the output shaft 17 of the motor 25 (and the motor cylinder block 34). In this condition, if this rotational drive power is transmitted through the axles 23R and 23L to the wheels, then the wheels are driven in the forward direction. The rotational speed of the motor output shaft 17 increases as the skew angle α of the pump swash plate increases. When the skew angle becomes the maximum skew angle in the forward drive direction α F(MAX), the condition becomes the one indicated by vertical line b in FIG. 2. Therefore, the total speed ratio e increases from zero (vertical line a) to e1 (vertical line b). On the other hand, the rotational speed of the pump input shaft 10 decreases while the rotational speed of the motor output shaft 17 increases because the mechanical power transmission through the small pinion 14, the large sun gear 16, the interdrive gear 18 and the interdriven gear 19 (i.e., through the power transmission mechanism) progresses at the same time.

After the pump swash plate achieves the maximum skew angle in the forward drive direction α F(MAX) (i.e., reaches the condition indicated by vertical line b), the motor swash plate is swiveled to make the skew angle β thereof gradually small from the maximum angle. As the skew angle becomes smaller and smaller, the rotational speed of the motor output shaft 17 increases further above the speed which is indicated by vertical line b. When the skew angle β becomes zero (i.e., the swash plate comes to the upright position), the rotational speed of the motor output shaft reaches the maximum speed (i.e., the total speed ratio becomes e2, and this condition is indicated by vertical line c in the figure).

However, as mentioned above, while the rotational speed of the motor output shaft 17 increases, the mechanical power transmission through the power transmission mechanism 4 increases, and the rotational speed of the pump input shaft 10 decreases. The gear ratio of the power distribution mechanism 3 and the power transmission mechanism 4 is predetermined in such a way that the rotational speed of the pump input shaft 10 (and the pump cylinder block 28) becomes zero when the skew angle β of the motor swash plate becomes zero (i.e., the swash plate comes to the upright position). Therefore, when the motor swash plate is at the upright position (β=0), the motor cylinder block 34 is freely rotatable, and the pump cylinder block 28 is hydraulically locked and becomes stationary. In this condition (which is indicated by vertical line c in the figure), only the mechanical power transmission through the power transmission mechanism 4 is performed.

On the other hand, if the pump swash plate 32 is swiveled counterclockwise, starting from the condition indicated by vertical line a in the figure, then the oil is discharged from the pump 24 in the direction opposite to that described above in the hydraulic closed circuit 26. This reversed flow of oil to the motor 25 drives the motor output shaft 17 (and the motor cylinder block 34) in the direction opposite to that described above (i.e., in the reverse drive direction). The rotational speed of the motor output shaft 17 increases as the skew angle α of the pump swash plate increases. When the skew angle becomes α R(MAX), the condition reaches the point which is indicated by vertical line d in FIG. 2. In this way, the total speed ratio e changes from zero (vertical line a) to a negative value (e3).

Hydraulic closed circuit for power transmission

Now, with reference to FIG. 3, the hydraulic closed circuit 26 and the hydraulic circuit control system of the hydrostatic unit 2 for continuous speed variation are described. In this figure, the hydraulic pump 24 and the hydraulic motor 25 are indicated in symbols. The hydraulic closed circuit 26 comprises a first oil passage 26 a which connects one port 24 b of the pump 24 with one port 25 a of the motor 25 and a second oil passage 26 b which connects the other port 24 a of the pump 24 with the other port 25 b of the motor 25.

As mentioned previously, the swash plate 32 of the pump 24 is swiveled in either direction, clockwise or counterclockwise, from the upright position (neutral position). When the pump swash plate is swiveled clockwise (i.e., in the forward drive direction), the oil sucked from the port 24 a is discharged from the port 24 b. This oil is supplied to the motor 25 through the port 25 a to drive the motor 25 in the forward drive direction. Then, the oil is discharged from the port 25 b, again to be sucked into the port 24 a, thus circulating in the closed circuit 26. In this condition, if the wheels are driven by the rotation of the motor 25, then the pressure in the first oil passage 26 a becomes high in correspondence, and the pressure in the second oil passage 26 b become low. On the other hand, if the rotational speed of the wheels are being reduced by engine brake in coasting, for example, then the pressure in the second oil passage 26 b becomes high in proportion to the force of engine brake, and the pressure in the first oil passage 26 a become low.

When the pump swash plate 32 is swiveled counterclockwise (i.e., in the rearward drive direction), the oil flow is reversed from that described above. This reversed oil flow drives the motor in the rearward drive direction. Likewise, the conditions of the pressure in the first and second oil passages 26 a and 26 b become opposite of what is described above.

In this way, the power transmission between the hydraulic pump and the hydraulic motor is performed. However, in this power transmission, while the oil is circulated in the hydraulic closed circuit 26, heat is generated, and the temperature of the oil rises. Also, in the circulation, contaminants may be collected in the oil, or some part of the oil may leak, for example, through the clearance of the plungers, into an oil tank. To solve such problems, the system enables part of the oil to be exchanged for cooling, supplementary feed, and purification (or flushing). A charge pump 43 is provided to supply oil in the oil tank 41 to a first line 100 through a suction filter 42. The charge pump 43 is driven by the engine E directly, so the amount of discharge is proportional to the rotation of the engine.

The oil discharged from the charge pump 43 into the first line 100 is regulated by a regulator valve 60 to have a predetermined line pressure PL. The first line 100 is branched as shown in the figure, and one branched first line 100 a is connected to a modulator valve 65, which includes a reducing valve. The output port of the modulator valve 65 is connected to a second line 101 to adjust the pressure of the second line to a predetermined modulated pressure Pm. FIG. 5 shows the construction of the modulator valve 65, which includes a spool 66, which is biased leftward by a spring 67, in a housing. The hydraulic pressure in the branched first line 100 a, which is connected to a port 65 a of the modulator valve 65, is reduced to a pressure (a constant pressure) at which the force of the spring 67 balances with the force of the pressure in the control line 101, thus creating the modulated pressure Pm in the control line 101. In the figure, mark “x” indicates a drainage line.

The second line 101 also branches into a plurality of lines, and one of the branched line, a branched second line 101 a, leads to a first linear solenoid valve 51. The first linear solenoid valve 51 adjusts the modulated pressure Pm to generate a control pressure PCL in proportion to a control current (I) as shown in FIG. 6, and the control pressure PCL acts on the regulator valve 60 through a control line 110.

FIG. 4 shows the construction of the regulator valve 60, which comprises a spool 61, which is slidable leftward and rightward, and a spring 62, which biases the spool 61 leftward, in a housing. The regulator valve has a plurality of ports 60 a˜60 e in the housing. The ports 60 a and 60 b are connected to the first line 100, the port 60 c is connected to a charge line 130, the port 60 d is connected to a discharge line 131, and the port 60 e is connected to the above mentioned control line 110.

In this condition the left end of the spool 61 of the regulator valve 60 is exposed to the line pressure PL from the first line 100 through the internal fluid-communication hole 61 a, and the right end is pushed by the biasing force of the spring 62 and exposed to the control pressure PCL. As mentioned above, the control pressure PCL is adjustable with the first linear solenoid valve 51, so the line pressure PL is controllable by controlling the control current (I), which is applied to the first linear solenoid valve 51, as shown in FIG. 7. In the control of the line pressure PL with the regulator valve 60, excess oil flows into the charge line 130 when the port 60 a comes into fluid communication with the port 60 c as the spool 61 shifts to the right. Further excess oil flows into the discharge line 131 when the port 60 b comes into fluid communication with the port 60 d.

As shown in FIG. 3, the charge line 130 is connected with a charge supply lines 105 a and 106 a, which are connected with the first and second oil passages 26 a and 26 b respectively through check valves 44 a and 44 b. The oil flowing into the charge line 130 must pass through either the check valve 44 a or the check valve 44 b before being supplied into either the first oil passage 26 a or the second oil passage 26 b whose pressure is lower than the other. In this way, the oil in the hydraulic closed circuit 26 is replenished.

The oil discharged into the discharge line 131 is cooled by an oil cooler 151, and then it is returned through a lubricator 152 into the tank 41.

As shown in FIG. 3, the first and second oil passage 26 a and 26 b constituting the hydraulic closed circuit 26 are connected with discharge lines 105 b and 106 b respectively, and these discharge lines 105 b and 106 b are connected to a shuttle valve 70. FIG. 8 shows the construction of the shuttle valve 70, which comprises a spool 71, which is slidable rightward and leftward in a housing, a pair of springs 72 and 73, one of which biases the spool 71 to the right, and the other to the left.

The hydraulic pressures of the discharge lines 105 b and 106 b work on the right and left ends of the spool 71 respectively. When the pressure of either one of the first or second oil passage 26 a or 26 b becomes higher than the other, the spool 71 is shifted by the pressure difference, and either the left side port 70 a or the right side port 70 b of the shuttle valve 70 whose hydraulic pressure is lower than the other is connected with the port 70 c which leads to a discharge line 132 to discharge some oil from the line which has the lower pressure. In this way, the oil is discharged in the amount which corresponds with the amount replenished in the hydraulic closed circuit 26, and the hydraulic oil is cooled and flushed, etc. In addition, the discharge line 132 is provided with a low pressure relief valve 74, which regulates the pressure of the line having the lower pressure. The oil discharged into the discharge line 132 is also cooled by the oil cooler 151 and returned through the lubricator 152 into the tank 41.

The hydraulic closed circuit 26 is also provided with high pressure relief valves 75F and 75R, which are connected through relief lines 105 c and 106 c to control the maximum pressure of the first oil passage 26 a and the second oil passage 26 b. These relief valves are identical in construction, so here, the high relief valve 75F, one of these two valves, is described with reference to FIG. 9.

The high relief valve 75F has two independent spools 76 and 77 in a housing. A first spool 76 is biased rightward by a first spring 79 a to cut off the fluid communication between the right side port 75 a and a port 75 b located inward in the housing, but if the spool is shifted leftward against this biasing force, then the two ports 75 a and 75 b are brought into fluid communication. The port 75 a is connected to the first oil passage 26 a through the relief line 105 c, and the port 75 b is connected to a branched charge line 130 a. In addition, an orifice 76 a is provided in the first spool 76, so the pressure in the first oil passage 26 a provided through the port 75 a acts on both the right and left sides of the first spool 76. Therefore, the first spool 76 is positioned at the right end of its stroke by the bias of the first spring 79 a in normal condition.

On the other hand, the second spool 77 is biased leftward by a second spring 79 b whose right end is also biasing a blocking valve piece 78 rightward in the housing. In this condition, the blocking valve piece 78 blocks a fluid communication passage 79 c which is provided with an orifice to communicate with the valve chamber of the first spool 76. The left side port 75 c, to which the left end of the second spool 77 faces, is connected through a control line 107 to a second linear solenoid valve 52, which adjusts the line pressure PL of a branched line 100 b in correspondence with a control current to provide a control pressure PCH to the control line 107. In this condition, the pressure on the left end of the second spool 77 can be controlled by controlling the control current applied to the second linear solenoid valve 52.

In the high pressure relief valve 75F, the pressure of the first oil passage 26 a acts on the blocking valve piece 78 from the right, and the force of the second spring 79 b act on the blocking valve piece 78 from the left. Because the left end of the second spring 79 b is placed on the second spool 77, the force of the control pressure PCH from the control line 107 acting on the second spool 77 is added to the rightward biasing force. Therefore, the biasing force acting on the blocking valve piece 78 rightward is controllable by adjusting the control pressure PCH, which is controlled by the second linear solenoid valve 52.

By this rightward biasing force, the blocking valve piece 78 blocks a fluid communication passage 79 c. However, if the pressure of the first oil passage 26 a rises, and the force from this pressure acting on the blocking valve piece 78 leftward exceeds the above rightward biasing force, then the blocking valve piece 78 is shifted leftward opening the fluid communication passage 79 c to the drainage. As a result, an oil flow is generated through the orifice 76 a of the first spool 76, which in turn creates a difference between the pressures on the right and left sides of the first spool 76. By this pressure difference, the first spool 76 is shifted leftward, and the port 75 a and the port 75 b are brought into fluid communication. As a result, some oil is discharged from the first oil passage 26 a to the branched charge line 130 a, and this discharged oil is fed into the second oil passage 26 b through the check valve 44 b.

In summary, when the pressure of the first oil passage 26 a rises above a predetermined value, the blocking valve piece 78 is released, and the first spool 76 shifts to the left, which discharges oil from the first oil passage 26 a to the second oil passage 26 b to maintain the pressure of the first oil passage 26 a at or below the predetermined value. On the other hand, if the pressure of the second oil passage 26 b rises above the predetermined value, the other high pressure relief valve 75R discharges oil from the second oil passage 26 b to the first oil passage 26 a to prevent the pressure of the second oil passage 26 b from rising above the predetermined value. In this way, the pressures of the first and second oil passages 26 a and 26 b are prevented from rising above the predetermined value by the high pressure relief valves 75F and 75R, which value or pressure is adjustable by controlling the current applied to the second and third linear solenoid valves 52 and 53 as described above. In this example, the high relief pressure (PHF and PHR) is adjustably set in proportion with the control current (I) as shown in FIG. 10.

In this embodiment, the control block 27 of the hydraulic pump 24 and the motor 25 is provided with respective valve plates, each having a configuration shown in FIG. 11, in the valve plate face 27 a. These valve plates, one provided for the pump and the other for the motor, may differ from each other in size, but their basic configurations and functions are identical. Thus, only the valve plate 150 of the pump 24 is described here with reference to FIG. 11.

In the pump 24, the pump cylinder block 28 is driven by the engine and is rotated in the direction indicated with arrow A in the figure (clockwise) in contact with the valve plate 150. When the pump swash plate 32 is swiveled in the forward drive direction, the pump plungers 30 reciprocate in the cylinder bores in synchronization with the rotation of the pump cylinder block 28. In this reciprocation, the pump plungers 30 reach the top dead center (T.D.C.) at the lower end of the valve plate as shown in FIG. 11, and they reach the bottom dead center (B.D.C.) at the upper end of the valve plate as the cylinder block rotates. While the pump plungers 30 travel from the top dead center (T.D.C.) to the bottom dead center (B.D.C.) above the left half of the valve plate as shown in the figure, oil is sucked, and while they travel from the bottom dead center (B.D.C.) to the top dead center (T.D.C.) above the right half of the valve plate, the oil is discharged.

Therefore, the valve plate 150 is provided with a semicircular first port 151 in the left half and with a semicircular second port 152 in the right half of the valve plate. The first port 151 is connected to the second oil passage 26 b, and the second port 152 is connected to the first oil passage 26 a. In addition, the valve plate is provided with main notches 151 a and 152 a, which extend to the respective entrances of the first and second ports 151 and 152 in the rotational direction. These main notches are to moderate the rapid pressure change which occurs when the cylinder bores 29 are brought into fluid communication with the ports 151 and 152 as the pump cylinder block 28 rotates. Furthermore, subnotches 153 and 154 are provided independently from and in parallel with the main notches 151 a and 152 a as shown in the figure.

As the subnotches 153 and 154 are independent and away from the main notches 151 a and 152 a and the ports 151 and 152, no notch effect can be expected in this condition. However, the subnotches 153 and 154 are connected to the inlets of the first and second ports 151 and 152 through shortcircuiting oil passages 155 a and 155 b and 156 a and 156 b as shown in dotted line in the figure, and variable notch valves 80A and 80B are provided on the shortcircuiting oil passages to open and close the shortcircuiting oil passages.

Both the variable notch valves 80A and 80B have an identical construction, which is shown in the FIG. 12 (this figure shows the variable notch valve 80A). The variable notch valve 80A comprises a valve spool 81, which is slidable rightward and leftward, a spring 82, which biases the valve spool 81 rightward, and a supporting spool 83, which slidably engages with the left end portion of the valve spool 81, in a housing. The housing is provided with a first port 80 a, which is connected to the shortcircuiting oil passage 155 a, a second port 80 b, which is connected to the shortcircuiting oil passage 155 b, and a third port 80 c, which is connected to a control line 102. When the valve spool 81 is biased by the force of the spring 82 and positioned at the right side as shown in the figure, the right end portion of the spool 81 cuts off the fluid communication between the first and second ports 80 a and 80 b. When the valve spool 81 is shifted to the left, theses ports are brought into fluid communication. The pressure in the first port 80 a is communicated through a bore 81 a which is provided in the spool 81 to the chamber 84 which exists between the valve spool 81 and the supporting spool 83. Therefore, this pressure will not provide any thrust to the valve spool 81.

As shown in FIG. 3, the control line 102 is connected with a branched second line 101 b through an orifice 45 a and to an open-close control solenoid valve 45, which can open the control line 102 to the drainage. When the open-close control solenoid valve 45 opens the control line 102 to the drainage, the pressure in the control line 102 is lowered. On the other hand, when the open-close control solenoid valve 45 closes or cuts off the fluid communication of the control line 102 to the drainage, the modulated pressure Pm is supplied from the branched second line 101 b to the control line 102.

When the control line 102 is maintained at the low pressure by the opening of the open-close control solenoid valve 45, the valve spool 81 of the variable notch valve 80A is positioned at the right side by the biasing force of the spring 82, blocking the shortcircuiting oil passages 155 a and 155 b. In this condition, the subnotch 153 has no effect on the system. On the other hand, when the control line 102 is supplied with the modulated pressure Pm by the closing of the open-close control solenoid valve 45, the valve spool 81 is shifted to the left side by this pressure, bringing the shortcircuiting oil passages 155 a and 155 b into fluid communication. In this condition, the subnotch 153 becomes effective.

This description is made of the variable notch valve 80A, but the same can be said of the other variable notch valve 80B, which is controllable in the same way as the variable notch valve 80A. Also, the hydraulic motor 25 is provided with such subnotches, and the pressure of a control line 103 is variably set with an open-close control solenoid valve 46 to control variable notch valves 80C and 80D, which control the subnotches.

Skew angle control system for the swash plates

Now, a control system for swiveling the pump swash plate 32 and the motor swash plate 38 is described with reference to FIG. 13. This control system uses the line pressure PL, which is supplied through a branched first line 100 d, and the modulated pressure Pm, which is supplied through a branched second line 101 c. In FIGS. 3 and 13, circled letters “a” and “b” in one of the figures are connected to the respective circled letters of the other figure.

For swiveling the pump swash plate 32, a pair of servo cylinders 92 a and 92 b are provided slidably in servo cylinder bores 91 a and 91 b as shown in the figure, and the servo cylinder bores 91 a and 91 b are connected through servo control lines 121 and 122 to a pump control valve 84, which is a four way valve. Depending on the position of the spool 85, the pump control valve 84 switches the supply of the line pressure PL from the branched first line 100 d either to the servo control line 121 or to the servo control line 122. The spool 85 is biased rightward by a spring 86, and the right end of the spool receives the pressure coming through the right end port 84 a. The balance of the forces from the spring and the pressure sets the position of the spool in the housing.

When the pressure of the right end port 84 a is controlled, the position of the spool 85 is adjusted to operate the servo cylinders 92 a and 92 b, which eventually regulates the pump swash plate 32. To control the pressure of the right end port 84 a, a pump control pressure PCP is supplied to the port 84 a from a fourth linear solenoid valve 54 through a control line 111. The fourth linear solenoid valve 54 controls the modulated pressure Pm of the branched second line 101 c to generate the pump control pressure PCP in proportion with a control current and leads the pump control pressure into the control line 111. Thus, by controlling the control current applied to the fourth linear solenoid valve 54, the pump control valve 84 is controlled to control the skew angle of the pump swash plate 32.

The control of the skew angle of the motor swash plate 38 is operated in the same manner. A pair of servo cylinders 96 a and 96 b are provided slidably in servo cylinder bores 95 a and 95 b, which are connected through servo control lines 123 and 124 to a motor control valve 87. In the same way as the pump control valve 84, the pressure of the right end port 87 a is controlled to control the position of the spool 88, which controls the servo cylinders 96 a and 96 b to regulate the motor swash plate 38. To control the pressure of the right end port, a motor control pressure PCM is supplied to the port 87 a from a fifth linear solenoid valve 55 through a control line 112. Thus, by controlling the control current applied to the fifth linear solenoid valve 55, the motor control valve 87 is controlled to control the skew angle of the motor swash plate 38.

In this embodiment, the hydraulic pump 24 is provided with a lock-up brake 93, which can hold the pump cylinder block 28 stationary. As previously mentioned, when the skew angle of the pump swash plate becomes the maximum skew angle in the forward drive direction, i.e., α=α F(MAX), and the skew angle of the motor swash plate becomes zero (β=0) in the forward drive condition (represented by vertical line c in FIG. 2), in theory, the rotational speed of the pump cylinder block 28 comes to zero, and the power transmission is carried out only through the power transmission mechanism 4 (if there is no transmission loss). However, in reality, there is a loss from frictional drag, oil leak, etc., which causes the pump cylinder block 28 to rotate a little and reduces the efficiency of the power transmission through the power transmission mechanism 4. To solve this problem, the pump cylinder block 28 is held stationary with the lock-up brake 93 to perform only the mechanical power transmission and to improve the transmission efficiency.

The lock-up brake 93 is a wet multiple disc brake, and it is operated in cooperation with a piston 104. The lock-up pressure PLB which empowers the piston 104 to squeeze the brake is generated by a sixth linear solenoid valve 56 and supplied through a lock-up line 113. The sixth linear solenoid valve 56 generates the lock-up pressure PLB, which is proportional to the control current supplied to the solenoid valve. Therefore, it is possible to control the lock-up brake 93 to engage partially or completely as much as desired.

Operational control

Now, the operational control of the variably continuous transmission T, which has the above construction and control circuit, is described. FIG. 14 shows the general construction of the control system. A control unit ECU generates control signals to the solenoid valves 45, 46 and 51 ˜56 (these valves constitute the hydraulic circuit, and they have been already explained) in response to detection signals which are received from various sensors 201˜214. The control unit ECU comprises an input interface to receive signals from the sensors and an output interface to send signals to the solenoid valves. The control unit ECU is operated by the central processing unit CPU, which operates on a software program that is stored in memory ROM 1 and ROM 2 to generate the control signals.

The sensors include an engine rotation sensor 201, which detects the rotational speed Ne of the engine E, a vehicle speed sensor 202, which detects the speed of the vehicle V, a shift position sensor 203, which detects the position of the shift lever, a brake sensor 204, which detects the operation of the brake of the vehicle, a throttle opening sensor 205, which detects the degree of the throttle opening θ TH or how much the engine throttle is opened, a pump swash plate angle sensor 206, which detects the skew angle α of the pump swash plate, a motor swash plate angle sensor 207, which detects the skew angle β of the motor swash plate, a pressure sensor 208, which detects the pressure P1 of the first oil passage 26 a that rises high when the vehicle is accelerated, a pressure sensor 209, which detects the pressure P2 of the second oil passage 26 b that rises high when the vehicle is decelerated, a creep lever sensor 210, which detects the amount of manipulation of the creep lever operated by the driver to set a creeping force, an EXC switch sensor 211, which detects the operation of an EXC switch that renews the initial value for the skew angle of the pump swash plate, a PB sensor 212, which detects the negative pressure of the engine in suction, a transmission oil temperature sensor 213, which detects the temperature of the oil in the transmission, and an engine water temperature sensor 214, which detects the temperature of the engine cooling water.

The operational control of the transmission follows a main control flow which is shown in FIG. 15. At first, at Step B1, an initialization control is executed to set an initial value to the output current which is supplied to the solenoid valves. Then, a target engine rotational speed Ne' is set on the basis of the vehicle speed V, the throttle opening θ TH and the shift lever position at Step B2. This step is executed, for example, by using a map such as shown in FIG. 16. The map shows the target engine rotational speed Ne' in ordinate in relation with the vehicle speed V in abscissa for respective throttle openings. When the vehicle speed V and the throttle opening θ TH at present are detected, a value for the target engine rotational speed Ne' is taken in correspondence from the map. The map shown in FIG. 16 is for the condition where the shift lever is positioned at the D range, and different maps are prepared for other conditions where the shift lever is positioned at other ranges. In this way, with the detection of the shift lever position, the appropriate map is selected in correspondence with the position of the shift lever to set the target engine rotational speed Ne'.

Then, the control flow proceeds to Step B3, and a determination is made to proceed either to an angle control (B4) or to an Ne control (B5) (the details of these controls are described later). The angle control (B4) is to set the skew angle of the pump swash plate to a predetermined angle. This will be described more in detail later. The Ne control is to control the skew angles of the pump swash plate and motor swash plate and the operation of the lock-up brake in feedback on the basis of the engine rotational speed Ne. This will be also described more in detail later.

While either the angle control (B4) or the Ne control (B5) is being executed, the high pressure relief valves 75F and 75R are controlled at Step B6; the operation of the variable notch valves 80A˜80D is controlled at Step B7; and the line pressure PL is set by the regulator valve 60 at Step B8. These steps are repeated as shown in FIG. 15, and these steps will be also described more in detail later.

Now, the determination performed at Step B3 is described with reference to FIG. 17. At first, a determination is made whether the shift position is in the N or P range or not at Step C1. If the shift position is in the N or P range, then the angle control B4 is performed. If the shift position is in any other range, then another determination is made whether the vehicle speed V is in acceleration or not, i.e., whether the speed is increasing or not at Step C2. If the vehicle is not in acceleration, then yet another determination is made whether the vehicle speed V is equal to or less than a predetermined speed V1 (e.g., 5.0 km/h) or not at Step C3. If it is equal to or less than the predetermined speed, then the angle control B4 is performed. However, if it is above the predetermined speed, then the control flow proceeds to Step C4.

On the other hand, if the vehicle is determined as being accelerating at Step C2, then the control flow proceeds to Step C6, and a determination is made of the throttle opening θ TH of the engine whether the throttle opening is about zero or not (i.e., whether the accelerator pedal is released or not). If the throttle opening θ TH is determined to be not zero, then the control flow proceeds to Step C4. However, if it is determined as almost zero, then the control flow proceeds to Step C7, where another determination is made whether the vehicle speed V is equal to or less than the predetermined speed. If it is so, then the angle control B4 is performed. The driver is judged as not in consideration of accelerating the vehicle if the throttle is closed completely and the vehicle is running very slowly (at the predetermined speed or less) even though the vehicle is in acceleration, so the angle control B4 is carried out. However, if the vehicle speed V is above the predetermined speed, then the control flow proceeds to Step C4.

At Step C4, a determination is made of the actual rotational speed Ne of the engine whether the actual rotational speed Ne is less than the target engine rotational speed Ne', which is set at Step B2. If the actual rotational speed Ne is determined to be greater or equal to the target rotational speed Ne', then the Ne control B5 is carried out. On the other hand, if the actual rotational speed Ne is determined to be less than the target rotational speed Ne', then another determination is made of the skew angle α of the pump swash plate and whether the angle α is equal to or smaller than a predetermined value α' or not at Step C5. If it is equal to or smaller than the predetermined value, then the Ne control B5 is carried out. However, if it is above the predetermined value, then the angle control B4 is carried out. For this predetermined value α' , the control flow uses the target skew angle α' for the swash plate that is determined in another control method determination flow, which will be described later with reference to FIGS. 18 and 19.

The above control flow for control method determination makes clear that the angle control B4 is performed in the following conditions:

(1) the position of the shift lever is at the N (neutral) or P (parking);

(2) the vehicle is not in acceleration, and it is driving at a speed which is equal to or less than the predetermined speed;

(3) the throttle is closed completely, and the vehicle speed is equal to or less than the predetermined speed even though the vehicle is in acceleration; and

(4) the skew angle α of the pump swash plate is equal to or smaller than the predetermined value, and the actual engine rotational speed Ne is less than the target engine rotational speed Ne'.

Now, the angle control B4 is described with reference to FIGS. 18 and 19. It is clear from the above mentioned control flow that the angle control B4 is to be performed when the transmission is close to its neutral condition (in the vicinity of vertical line a in FIG. 2). The skew angle β of the motor swash plate is set either at zero or at the maximum value β (MAX). This means that when the shift position is in the N or P range, the skew angle β is set at zero, and the skew angle β is set at the maximum value β (MAX) in other cases. The skew angle of the motor swash plate is easily set at these angles by adjusting the motor control pressure PCM, which is supplied to the right end port 87 a of the motor control valve 87, to zero or to the maximum value so as to shift the spool 88 to the left end position or to the right end position.

In addition to the above control of the skew angle of the motor swash plate, the angle control B4 is to carry out the positional feedback control of the skew angle of the pump swash plate. In this control, first, a determination is made whether the shift position is in the N or P range or not at Step D1. If the shift position is in the N or P range, then the target deviation α 1 of the pump swash plate is set at zero (α 1=0) at Step D6.

If the shift position is in any range other than the N or P range, then a determination is made of the engine rotational speed Ne whether the rotational speed Ne is equal to or less than a predetermined rotational speed (e.g., 600 rpm, which is a little less than an idling rotational speed, i.e., 700 rpm) at Step D2. If the engine rotational speed Ne is equal to or less than the predetermined rotational speed, then the target deviation α 1 of the pump swash plate is set at zero (α 1=0) at Step D6. From this control, it is understood that when the positional feedback control for the pump swash plate (angle control) is performed, the target deviation α 1 is set to zero (α 1=0) unconditionally if the engine rotational speed is equal to or less than the idling rotational speed so as to prohibit any control which may increase the delivery of the pump.

On the other hand, if the engine rotational speed is above the predetermined value, then a determination is made whether the engine throttle opening is completely closed (θ TH=0) or not at Step D3. If the throttle is open, then the control flow proceeds to Step D8, where the target deviation α 1 of the pump swash plate is set. If the throttle is completely closed, then the control flow proceeds to Step D4, and a determination is made whether the vehicle speed V is equal to or less than a predetermined vehicle speed (e.g., 5 km/h) or not. If the vehicle speed V is above the predetermined speed, then the control flow proceeds to Step D8. However, if it is equal to or less than the predetermined vehicle speed, then the control flow proceeds to Step D5, where another determination is made whether the brake operation is ON or not, i.e., whether the brake pedal is being applied or not. If the brake is ON, then the control flow proceeds to Step D6. If it is OFF, then the control flow proceeds to Step D8.

Since the target deviation α 1 of the pump swash plate is set at Step D6 or D8, the setting carried out at Step D8 is described at first. As shown in FIG. 20, a creep lever 220 and an EXC switch 221 are provided at the driver seat. The creep lever 220 is manipulated by the driver, and the EXC switch 221 is used to renew the initial value of the skew angle of the pump swash plate. At Step D8, the target deviation α 1 of the pump swash plate is set on the basis of the manipulated amount ACR of the creep lever 220. Specifically, a relation between the manipulated amount ACR of the creep lever 220 and the target deviation α 1 of the pump swash plate is predetermined as shown in FIG. 21, and the setting is performed in accordance with this relation.

As shown in FIG. 20, the creep lever 220 is swiveled forward and backward from an upright position, which is a neutral position N. The creep lever 220 is an automatically resetting lever which automatically returns to the neutral position, so when it is not manipulated, it is positioned at the neutral position N as shown in the figure. When the creep lever 220 is at the neutral position N, the target deviation α 1 of the pump swash plate is zero. If the lever 220 is swiveled to the “+” side as shown in the figure, then the target deviation α 1 is set to a positive value which is proportional to the manipulated amount of the creep lever. Likewise, if the creep lever is swiveled to the “−” side, then the target deviation α 1 is set to a negative value which is proportional to the manipulated amount of the creep lever.

At Step D11, the initial value α 0 of the skew angle of the pump swash plate and the initial value β 0 of the skew angle of the motor swash plate are set. These values are set on the basis of the shift position as shown in FIG. 22. If the shift position is in the N, P or R range, then these values are set to respective predetermined values notwithstanding the condition, i.e., ON or OFF, of the brake. However, if the shift position is in any forward drive range (D, L, S or M range), then these values are set at respective different predetermined values which are dependent on the condition, i.e., ON or OFF, of the brake. In this case, as the brake is OFF, Predetermined value 2 is selected. In the table of the figure, Predetermined value 1 which is selected for the R (rearward drive) range is a negative value (e.g., −5 degrees), and the predetermined values 2 and 3 which are selected for the D, L, S and M (forward drive) ranges are positive values (e.g., +5 degrees and +3 degrees, respectively).

From the target deviation α 1 of the pump swash plate and the initial value α 0 of the skew angle, a target skew angle α' is calculated for the pump swash plate at Step D12. This calculation is described with reference to FIG. 23. As mentioned previously, the target deviation α 1 of the pump swash plate is set when the creep lever 220 is manipulated by the driver. More specifically, the target deviation α 1 of the pump swash plate is taken for the calculation of the target skew angle α' when the driver presses the EXC switch 221 down while he is manipulating the creep lever (i.e., when the driver shows his intention to set the value). The EXC switch 221 is a momentary switch, which is ON only while it is pressed down, and it returns to OFF when released.

At this step, a determination is made of the flag which is set up when the EXC switch is turned on whether the flag is one (EXC switch=1) or not. If the flag is zero (0), then the control flow proceeds to Step D54, and another determination is made of the target deviation α 1 of the pump swash plate whether the target deviation is zero (α 1=0) or not. If the target deviation is not zero, i.e., the creep lever 220 is being manipulated, then the control flow proceeds to Step D55, where the target skew angle α' is calculated (α'=α0+α1) At this point, a timer is set to count a predetermined time period at Step D56, and a determination is made of the condition of the EXC switch 221 whether the EXC switch is turned on or not at Step D57. If the switch is OFF, then the control flow return to the start of this control flow. However, if the switch is turned ON, then the flag is set up (EXC switch=1) at Step D58, and the target skew angle α' at this moment is memorized as α ex at Step D59.

When the flag is one (EXC switch=1), the control flow returns to the start of this control flow and proceeds to Step D51 and to Step D52, and the target skew angle (α '=α ex) is retained until the time which is set on the timer elapses. When the time is up, a determination is made of the target deviational a1 of the pump swash plate whether the target deviation is zero (α 1=0) or not at Step D54. If it is zero (α1=0), i.e., the creep lever 220 is released after the EXC switch 221 is pressed, the control flow proceeds to Step D60. There, another determination is made whether the flag is up (EXC switch=1) or not. If the flag is up (1), then the initial value α 0 for the skew angle of the pump swash plate is reset at Step D61 with the target skew angle α ex which is memorized at Step D59, and the flag is lowered (EXC switch=0) at Step D62. This reset initial value α 0 is then set as the target skew angle α' at Step D63.

Summarizing the above control at Step D12, the following can be said. If the EXC switch 221 is pressed down while the creep lever 220 is manipulated, then the target skew angle α' of the pump swash plate is set by adding the initial value α 0 of the skew angle of the pump swash plate to the target deviation α 1 which corresponds with the amount of manipulation of the creep lever at the moment. Then, this target skew angle α' is retained even though the creep lever 220 is returned to the neutral position. In addition, this target skew angle α' (=α ex) becomes a new initial value α 0 for the skew angle of the pump swash plate. Thereafter, if the EXC switch 221 is pressed down while the creep lever 220 is manipulated again, then the current target skew angle α' of the pump swash plate is added with the target deviation α 1 which corresponds with the amount of manipulation of the creep lever made at this moment. Thus, if the manipulation of the creep lever 220 and the pressing of the EXC switch 221 are repeated, then the target skew angle α' is set accordingly with the addition of the target deviations α 1 which correspond to the repeated operations.

After the target skew angle α' of the pump swash plate is set at Step D12 as described above, a determination is made at Step D13 whether this value α' is greater than the initial value α 0, which is set at Step D11. If the target skew angle α' is equal to or smaller than the initial value α 0 (α'≦α 0), then the target skew angle α' is set equal to the initial value (α '=α 0) at Step D14. When the creep lever 220 is manipulated to the negative side, the target skew angle α' may become smaller than the initial value (α'≦α 0). However, even if such case happens, the skew angle of the pump swash plate will not be set to a value smaller than the initial value α 0.

On the other hand, the target deviation α 1 of the pump swash plate is set to zero (α 1=0) at Step D6, and the initial value α 0 for the pump swash plate and the initial value β 0 for the motor swash plate are set in accordance with the table of FIG. 22 at Step D9. In this case, as the brake is ON, Predetermined value 3 is taken as an initial value when the shift position is in a forward drive range (D, L, S or M). Predetermined value 3 is smaller than Predetermined value 2 so as to make the initial value α 0 relatively small when the brake pedal is applied. This initial value α 0 and the target deviation α 1 are added to achieve the target skew angle α' for the pump swash plate at Step D10.

Then, a calculation is made for the deviation angle Δ α (=α−α ') of this target skew angle α' from the current skew angle a of the pump swash plate at Step D15, and a target current (target deviation current DICMDp) which is applied to the fourth linear solenoid valve 54 is set to realize this deviation angle in the pump swash plate control at Step D16. Relations between the target deviation current DICMDp and the deviation angle Δ α are predetermined on the basis of the operational characteristics of the pump swash plate 32, which correspond with the operation of the fourth linear solenoid valve 54 as shown in FIG. 24. The target deviation current DICMDp is taken from the figure in correspondence with the deviation angle Δ α which is calculated at Step D15.

With this target deviation current DICMDp, a feedback control is performed to set the skew angle of the pump swash plate to the target skew angle α'. At first, for deciding the direction for the control, a determination is made whether the current skew angle α is greater than the target skew angle α' or not at Step D17. If the current skew angle is greater than the target skew angle (α>α'), then a target current ICMDp is calculated at Step D19 by adding the target deviation current DICMDp to a control current IOp which is needed to retain the spool 85 of the pump control valve 84 at the neutral position. However, if the current skew angle is equal to or smaller than the target skew angle (α≦α'), then the target current ICMDp is calculated at Step D18 by subtracting the target deviation current DICMDp from the control current IOP which is needed to retain the spool 85 of the pump control valve 84 at the neutral position. With this target current ICMDp, the fourth linear solenoid valve 54 is actuated at Step D20 to control the skew angle of the pump swash plate closer to the target skew angle α'.

It is understood from the above description that when the vehicle is in the N or P range, the skew angle of the pump swash plate is set to zero, but when the vehicle is in any other drive range, the skew angle of the pump swash plate is controlled to the target skew angle α', and the vehicle comes into creeping condition. As the target skew angle α' changes with the amount of manipulation of the creep lever 220, the inching of the vehicle is possible with the operation of the creep lever.

As for the skew angle β of the motor swash plate, the initial value β 0 has been already set as mentioned previously. As this initial value is either zero or the maximum value, a maximum current or zero current is set as a target current ICMDm at Step D21. With this target current, the fifth linear solenoid valve 55 is actuated at Step D22 to control the motor swash plate to the upright position or to the maximally slanted position.

Here, the initial value α 0 for the skew angle of the pump swash plate which is set in the above mentioned angle control B4 is selected as shown in the table of FIG. 22. This value is set to prevent the vehicle which is stopped on a slope from involuntarily moving backward. As mentioned previously, when the shift position is in a forward drive range (D, L, S or M), Predetermined value 2 or Predetermined value 3, which gives a positive value to the skew angle of the pump swash plate, is selected. In this condition, as the skew angle of the motor swash plate is at the maximum, if the pump is driven by the engine which is in idling, for example, then the motor is driven in the forward drive direction because the oil discharged from the pump is supplied to the motor. Therefore, Predetermined value 2 is equivalent to the first pump capacity (especially the first pump capacity for a forward drive), and the skew angle of the motor swash plate at the maximum is equivalent to the predetermined motor capacity which is defined in the section for claims.

When the vehicle is stopped on an uphill road, the motor receives a driving force in the rearward drive direction. If there is oil leak in the hydraulic closed circuit, it might be expected that the motor would rotate at the rotational speed which corresponds with the speed of the leak, and the vehicle would move backward. In this system, oil discharged from the pump supplements the amount lost in an internal oil leak, so the motor is prevented from rotating to avoid any backward movement of the vehicle. If the brake pedal is applied, there will be no such problem, so the skew angle of the pump swash plate can be set to zero. However, the skew angle of the pump swash plate is set to Predetermined value 3, which is a positive value smaller than Predetermined value 2, in this embodiment to realize smooth start of the vehicle. Therefore, Predetermined value 3 is equivalent to the second pump capacity which is defined in the section for claims.

The above description is made for the condition when the shift position is in a forward drive range. A similar control to this control is performed for the rearward (R) range. However, in this case, a negative value, Predetermined value 1, is selected, and the oil discharged from the pump works to drive the motor in the rearward drive direction. In this condition, if the vehicle is stopped on a downhill road, the motor receives a driving force in the forward drive direction. Even if there is oil leak in the hydraulic closed circuit, the oil discharged from the motor in the rearward drive direction supplements the deficiency created in an internal oil leak, so the motor is prevented from rotating to avoid any forward movement of the vehicle (i.e., the movement opposite to the direction set by the shift position). Therefore, Predetermined value 1 is equivalent to the first pump capacity for rearward drive.

In this embodiment, the skew angle of the pump swash plate is set to Predetermined value 1 notwithstanding the operation of the brake when the shift position is in the rearward range. However, the skew angle of the pump swash plate may be set to a value smaller than Predetermined value 1 as in the

case for a forward drive range with the brake operation.

Now, the Ne control B5 is described with reference to FIG. 25. This is to control the skew angles of the swash plates of the pump and the motor and the operation of the lock-up brake, thereby regulating the engine rotational speed Ne to the target engine rotational speed Ne' in feedback control. At first, a deviation rotational speed Δ Ne is calculated by subtracting the actual engine rotational speed Ne from the target engine rotational speed Ne' at Step E1 (Δ Ne=Ne'−Ne).

Then, at Step E2, a determination is made which quadrant the condition of the transmission is in, of the graph of FIG. 2. As mentioned previously, this graph describes the relation of the changes in the total speed ratio to the changes in the skew angles α and β of the swash plates of the pump and the motor. For controlling the speed change, this graph representing the condition of the transmission in control is divided into four control quadrants or zones (Quadrant I-IV) as shown in FIG. 26, and a determination is made which control quadrant the actual condition is in at Step E2. This determination is described in the following with reference to FIG. 27.

As shown in FIG. 27, first a determination is made at Step F1 whether the skew angle α of the pump swash plate is greater than an intermediate angle a m, which is between zero and the maximum value. This is to determine whether the skew angle α of the pump swash plate is near the maximum skew angle in the forward drive direction α F(MAX). If the skew angle α of the pump swash plate is equal to or smaller than the intermediate angle α m (α>α m), then the transmission is in the condition which corresponds to Quadrant I in FIG. 26, so the control flow proceeds to Step F5, whose control is described in detail in FIG. 28 (a). There, the control of the lock-up brake is turned OFF at Step F51; the control of the pump swash plate is turned ON at Step F52; and the control of the motor swash plate is turned OFF at Step F53. In this way, only the control of the pump swash plate is carried out for the condition in Quadrant I. Also, in this zone, the angle control B4 and the Ne control are switched, i.e., the controls for the start and stop of the vehicle are smoothly exchanged.

If the skew angle a of the pump swash plate is greater than the intermediate angle α m (α>α m), then a determination is made at Step F2 whether the skew angle β of the motor swash plate is greater than an intermediate angle β m, which is between zero and the maximum value. This is to determine whether the skew angle of the motor swash plate is near the maximum skew angle or not. If the skew angle of the motor swash plate is greater than an intermediate angle (β>β m), then the transmission is in the condition which corresponds to Quadrant II in FIG. 26, so the control flow proceeds to Step F6, whose control is described in detail in FIG. 28 (b). There, the control of the lock-up brake is turned OFF at Step F61; the control of the pump swash plate is turned ON at Step F62; and the control of the motor swash plate is turned ON at Step F63. In this way, the controls of the pump and motor swash plates are carried out at the same time for the condition in Quadrant II, and these two controls are carried over from one control to the other smoothly.

If the skew angle of the motor swash plate is equal to and smaller than an intermediate angle (β≦β m), then a determination is made whether the skew angle β of the motor swash plate is zero or not at Step F3. If the skew angle β of the motor swash plate is not zero, then the transmission is in the condition which corresponds to Quadrant III in FIG. 26, so the control flow proceeds to Step F7, whose control is described in detail in FIG. 28 (c). There, the control of the lock-up brake is turned OFF at Step F71; the control of the pump swash plate is turned OFF at Step F72; and the control of the motor swash plate is turned ON at Step F73. In this way, only the control of the motor swash plate is carried out for the condition in Quadrant III, and most part of the vehicle drive control is performed in Quadrant III and Quadrant IV, which is described in the following.

If the skew angle β of the motor swash plate is zero, then a determination is made whether the vehicle speed is equal to or less than a predetermined speed or not at Step F4. If the vehicle speed is equal to or less than the predetermined speed, then the above mentioned control for Quadrant III is performed at Step F7. In this way, the lock up is not carried out for the condition where the vehicle has a speed equal to or less than the predetermined speed. However, if the vehicle speed is greater than the predetermined value, then the transmission is in the condition which corresponds to Quadrant IV in FIG. 26, so the control flow proceeds to Step F8. There, the control of the lock-up brake is turned ON at Step F81, and a determination is made of the target current ICMDL which controls the sixth linear solenoid valve 56 that actuates the lock-up brake whether this current is equal to or smaller than a predetermined value or not at Step F82. If it is equal to or smaller than the predetermined value (i.e., the engaging force of the lock-up brake is relatively small), then the control of the pump is turned OFF at Step F83; and the control of the motor is turned ON at Step F84. However, if the target current is greater than the predetermined value (i.e., the lock-up brake is engaged to a certain degree), then not only the control of the pump is turned OFF at Step F72, but also the control of the motor is turned OFF at Step F73.

After the determination of the control zone is performed at Step E2 of FIG. 25 in this way, a gain coefficient K for speed change speed is set at Step E3. As the gain coefficient K is predetermined in relation to the vehicle speed V as shown in FIG. 29, the gain coefficient K is taken in correspondence with the actual vehicle speed V at the moment. It is clear from the figure that the smaller the vehicle speed V the greater the gain coefficient. Then, this gain coefficient K is multiplied by the deviation rotational speed Δ Ne, which is calculated at Step E1, to produce a speed change speed DI (=K×Δ Ne) at Step E4.

Thereafter, determinations are made whether the controls of the pump, the motor and the lock-up brake are ON or not at Steps E5, E7 and E9. If the control of the pump is ON, then the pump swash plate is controlled at Step E6; if the control of the motor is ON, then the motor swash plate is controlled at Step E8; and if the control of the lock-up brake is ON, then the lock up is controlled at Step E10.

First, the control of the pump swash plate executed at Step E6 is described with reference to FIG. 30. The pump swash plate is controlled to swivel in the forward drive direction α F or in the rearward drive direction α R. Although these directions are opposite to each other, the pump swash plate 32 is controlled in either direction in the same manner. Therefore, the control in the forward drive direction is described here, as an example. This is to control the current applied to the fourth linear solenoid valve 54, taking the engine rotational speed as the value to be controlled. For example, if the actual engine rotational speed Ne is less than the target engine rotational speed Ne', then the speed ratio is controlled to become greater (i.e., shifted toward LOW), which is to make the skew angle a of the pump swash plate smaller. On the other hand, if the actual engine rotational speed Ne is greater than the target engine rotational speed Ne', then the speed ratio is controlled to become smaller (i.e., shifted toward TOP), which is to make the skew angle α of the pump swash plate greater.

In this control flow, a target deviation current DICMDp is set at Step G1 as shown in FIG. 30. The target deviation current DICMDp is predetermined in relation with the speed change speed DI as shown in FIG. 31, and from this relation, a value which corresponds with the speed change speed DI(1) that is calculated at Step E4 in FIG. 25 is taken as the target deviation current DICMDp(1). Then, a determination is made whether the target engine rotational speed Ne' is greater than the actual engine rotational speed (i.e., the current engine rotational speed) Ne at Step G2.

If the target engine rotational speed is greater than the actual engine rotational speed (Ne'>Ne), then the control proceeds to Step G3, where a calculation is made for the target current ICMDp which is required for making the speed change ratio greater (i.e., to shift toward LOW). This target current is calculated by adding the target deviation current DICMDp to the control current IOp which is required for retaining the spool 85 of the pump control valve 84 at the neutral position. On the other hand, if the target engine rotational speed is equal to or less than the actual engine rotational speed (Ne'≦Ne), then the control proceeds to Step G4, where a calculation is made for the target current ICMDp which is required for making the speed change ratio smaller (i.e., to shift toward TOP). This target current is calculated by subtracting the target deviation current DICMDp from the control current IOp which is required for retaining the spool 85 at the neutral position.

With this target current ICMDp as a control current, the fourth linear solenoid valve 54 is actuated at Step G5. FIG. 32 shows the relation between the control pressure PC and the target current ICMDp which is applied to the linear solenoid valve 54. If the target current ICMDp equals the above mentioned control current IOp, then the pump control valve 84 is kept at the neutral. When the target current ICMDp is increased or decreased from this value, the speed is shifted toward LOW or toward TOP, respectively. In this way, the feedback control is performed to bring the actual engine rotational speed Ne close to the target engine rotational speed Ne'.

Now, the control of the motor swash plate which is carried out at Step E8 is described with reference to FIG. 33. This is to control the current applied to the fifth linear solenoid valve 55, taking the engine rotational speed as the value to be controlled. For example, if the actual engine rotational speed Ne is less than the target engine rotational speed Ne', then the speed ratio is controlled to become greater (i.e., shifted toward LOW), which is to make the skew angle β of the motor swash plate greater. On the other hand, if the actual engine rotational speed Ne is greater than the target engine rotational speed Ne', then the speed ratio is controlled to become smaller (i.e., shifted toward TOP), which is to make the skew angle β of the motor swash plate smaller.

In this control flow, a target deviation current DICMDm is set at Step H1 as shown in FIG. 33. The target deviation current DICMDm is predetermined in relation with the speed change speed DI as shown in FIG. 34, and from this relation, a value which corresponds with the speed change speed DI(1) that is calculated at Step E4 in FIG. 25 is taken as the target deviation current DICMDm(1). Then, a determination is made whether the target engine rotational speed Ne' is greater than the actual engine rotational speed (i.e., the current engine rotational speed) Ne at Step H2.

If the target engine rotational speed is greater than the actual engine rotational speed (Ne'>Ne), then the control proceeds to Step H3, where a calculation is made for the target current ICMDm which is required for making the speed change ratio greater (i.e., to shift toward LOW). This target current is calculated by adding the target deviation current DICMDm to the control current IOm which is required for retaining the spool 88 of the motor control valve 87 at the neutral position. On the other hand, if the target engine rotational speed is equal to or less than the actual engine rotational speed (Ne'≦Ne), then the control proceeds to Step H4, where a calculation is made for the target current ICMDm which is required for making the speed change ratio smaller (i.e., to shift toward TOP). This target current is calculated by subtracting the target deviation current DICMDm from the control current IOm which is required for retaining the spool 88 at the neutral position.

With this target current ICMDm as a control current, the fifth linear solenoid valve 55 is actuated at Step H5. FIG. 35 shows the relation between the control pressure MC and the target current ICMDm. If the target current ICMDm equals the above mentioned control current IOm, then the motor control valve 87 is kept neutral. As the target current ICMDm is increased or decreased from this value, the speed is shifted toward LOW or toward TOP, respectively. In this way, the feedback control is performed to bring the actual engine rotational speed Ne close to the target engine rotational speed Ne'.

Now, the control of the lock-up brake which is carried out at Step E10 is described with reference to FIG. 36. This is to control the current applied to the sixth linear solenoid valve 56 by determining the strength of the lock up from the ratio of the actual engine rotational speed Ne to the target engine rotational speed Ne'. For example, if the actual engine rotational speed Ne is less than the target engine rotational speed Ne', then the control current applied to the sixth linear solenoid valve 56 is decreased to make the lock up force smaller. On the other hand, if the actual engine rotational speed Ne is greater than the target engine rotational speed Ne', then the control current applied to the sixth linear solenoid valve 56 is increased to make the lock up force greater.

In this control flow, a target deviation current DICMDL is set at Step I1 as shown in FIG. 36. The target deviation current DICMDL is predetermined in relation with the speed change speed DI as shown in FIG. 37, and from this relation, a value which corresponds with the speed change speed DI(1) that is calculated at Step E4 in FIG. 25 is taken as the target deviation current DICMDL(1). Then, a determination is made whether the target engine rotational speed Ne' is greater than the actual engine rotational speed (i.e., the current engine rotational speed) Ne at Step I2.

If the target engine rotational speed is greater than the actual engine rotational speed (Ne'>Ne), then the control proceeds to Step 14, where a calculation is made for a new control current, i.e., a target current ICMDL which will be applied to the sixth linear solenoid valve 56 by subtracting the target deviation current DICMDL from the control current ICMDL' that is applied presently to the six linear solenoid valve 56. Then, this target current is set to weaken the present lock-up force. On the other hand, if the target engine rotational speed is equal to or less than the actual engine rotational speed (Ne'≦Ne), then the control proceeds to Step I3, where a calculation is made for a new control current, i.e., a target current ICMDL which will be applied to the sixth linear solenoid valve 56, by adding the target deviation current DICMDL' to the control current ICMDL that is applied presently to the sixth linear solenoid valve 56. Then, this target current is set to strengthen the present lock-up force.

With this target current ICMDL as a control current, the sixth linear solenoid valve 56 is actuated at Step I5. FIG. 38 shows the relation between the control pressure LC and the target current ICMDL, in which the control pressure LC is proportional to the target current ICMDL.

Now, the relief control at Step B6 of FIG. 15 is described. This control is to adjust the relief pressure that is applied to the high pressure relief valves 75F and 75R, which are provided on the first and second oil passages 26 a and 26 b comprising the hydraulic closed circuit 26. The pressures of the first and second oil passages 26 a and 26 b change in correspondence with the change of the vehicle drive direction (forward drive and rearward drive) and the condition of the vehicle speed (in acceleration or deceleration) as shown in FIG. 41. Therefore, this relief control responds to the change of these conditions to perform high pressure relief control. The pressure of the oil passage which is at the lower pressure is supplied from the low pressure relief valve 74 through the shuttle valve 70.

In this system, the pressure relief of the oil passage which is at the higher pressure is performed by the high pressure relief valves, which release quickly any overpressure and lower the pressure to a predetermined value, to alleviate the shock of surge pressure. As the maximum value for the high pressure differs depending on the skew angle of the pump swash plate when the vehicle is in acceleration, the target relief pressure is controlled in correspondence with the skew angle of the pump swash plate. Also, the target relief pressure is controlled in correspondence with the vehicle speed so that an appropriate engine brake force is realized while the vehicle is in deceleration.

The flow of this control is shown in FIGS. 39 and 40, and the circled letters, “B”, in the figures combine the two into one continuous flow. In this control, at first, a determination is made of the shift position whether it is in the N or P range at Step J1. If it is in the N or P range, then the control flow proceeds to Steps J8 and J9. However, if the shift position is in any other range than the N or P range, then a determination is made whether the engine rotational speed Ne is equal to or less than a predetermined rotational speed at Step J2. If it is equal to or less than the predetermined rotational speed, then the control flow also proceeds to Steps J8 and J9. However, if the engine rotational speed Ne is above the predetermined rotational speed, then a determination is made whether the vehicle speed V is equal to or less than a predetermined vehicle speed at Step J3. If it is above the predetermined vehicle speed, then the control flow proceeds to Steps J6 and J7. However, if the vehicle speed is equal to or less than the predetermined vehicle speed, then a determination is made of the engine throttle opening θ TH whether it is equal to or smaller than a predetermined throttle opening at Step J4. If it is above the predetermined throttle opening, then the control flow also proceeds to Steps J6 and J7. However, if the engine throttle opening is smaller than the predetermined throttle opening, then a determination is made whether the brake is ON or not at Step J5. If it is ON, then the control flow proceeds to Steps J8 and J9. However, if the brake is OFF, then the control flow proceeds to Steps J6 and J7.

A high relief pressure Pha is set in correspondence with the skew angle of the pump swash plate a as shown in FIG. 42 at Step J6. In the figure, the hydraulic pressure indicated by Line a1 is the maximum allowable pressure of the device, and the hydraulic pressure indicated by Line a2 is the target high pressure which alleviates surge pressure in correspondence with the skew angle of the swash plate. Furthermore, the high relief pressure Phb is set in correspondence with the vehicle speed V as shown in FIG. 43 at Step J7. This relief pressure is a target high relief pressure which is set on the basis of the required engine brake force.

On the other hand, a high relief pressure is set on the basis of the specific conditions which are determined at Steps J1-J5 as shown in FIG. 44 at Steps J8 and J9. Specifically, (1) if the shift position is in the N or P range (which is determined at Step J1), then the target relief pressure Ph is set to zero so that the system is made to bypass the closed circuit to establish the neutral condition. (2) if the engine rotational speed decreases below the idling rotational speed (i.e., a predetermined speed) (determined at Step J2), then the target relief pressure is set lower by the amount which corresponds to the rotational speed decrease to prevent the engine from stopping. (3) if the shift position is in any drive range other than the N or P range, the engine rotational speed is above the idling rotational speed, the vehicle speed is equal to or less than the predetermined speed (a relatively low speed), the throttle opening is equal to or smaller than the predetermined throttle opening (a relatively small opening), and the brake is operated, then the condition of the vehicle is judged as being equivalent to stopping, and the target relief pressure is set to the predetermined value. This predetermined value for the target relief pressure is set to zero for both acceleration and deceleration (Pha=0 and Phb=0), or it may be set to a small value which will give a smooth change when the driver releases the brake pedal and steps down the accelerator pedal.

As mentioned previously, the high relief pressure is adjusted as desired by controlling the currents applied to the second and third linear solenoid valves 52 and 53. After the target relief pressure is determined, the target control currents ICMDa and ICMDb which are necessary to set this relief pressure for acceleration or for deceleration are set at Steps J10 and J11. The relation between the target control current and the target relief pressure is shown in FIG. 45, and from this relation, the target control current ICMD which gives the target relief pressure Ph is taken.

Then, a determination is made whether the shift position is in the R (rearward) range or not at Step J12. If the shift position is not in the R range, then the second linear solenoid valve 52, which controls the relief pressure of the first oil passage 26 a, is actuated with the target control current ICMDa for acceleration at Step J13, and the third linear solenoid valve 53, which controls the relief pressure of the second oil passage 26 b, is actuated with the target control current ICMDb for deceleration at Step J14. However, if the shift position is in the R range, then the second linear solenoid valve 52 is actuated with the target control current ICMDb for deceleration at Step J15, and the third linear solenoid valve 53 is actuated with the target control current ICMDa at Step J16.

Now, the variable notch control B7 is described with reference to FIG. 46. This is to optimize the pressure change of each cylinder bore in correspondence with the discharges of the pump 24 and the motor 25 and the high pressure generated in the operation thereof. For example, an open and close control which is shown in the table of FIG. 47 is performed for the operation of the pump 24.

In this control flow, at first, the displacement volume Vp of the pump is determined from the skew angle α of the pump swash plate at Step K1. FIG. 48 shows the relation between the skew angle of the pump swash plate and the displacement volume of the pump, and it is calculated by Equation(2). Then the discharge of the pump Qp (=Vp×Np) is calculated at Step K2. Here, Np is the rotational speed of the pump, and it is calculated by Equation (3) with the engine rotational speed Ne and the rotational speed of the wheels Nv, which values are detected with the respective sensors.

Vp=S×PD×tan α×N  (2)

Here, S is the sectional area of the cylinder bore; PD is the diameter of the pitch circle of the cylinders; and N is the number of cylinders.

Np={Ne−i1×i2×(1+ip)×Nv}/ip  (3)

Here, i1 is the reduction ratio of the power transmission mechanism 4; i2 is the reduction ratio of the final speed reduction mechanism 5; and ip is the reduction ratio of the power distribution mechanism 3 (refer to FIG. 49).

Then, a determination is made whether the discharge of the pump Qp is greater than a predetermined value Qp0 or not at Step K3. If the discharge is greater than the predetermined value (Qp>Qp0), then the control flow proceeds to Steps K7 and K8. However, if the discharge is equal to or less than the predetermined value (Qp≦Qp0), then another determination is made of the high pressure Ph generated in the oil passage which is at the higher pressure whether this high pressure is greater than a predetermined pressure Ph0 or not at Step K4. If it is so (Ph>Ph0), then the control flow proceeds to Steps K5 and K6. However, if the high pressure Ph is equal to or smaller than the predetermined pressure Ph0 (Ph<Ph0), then the control flow proceeds to Steps K7 and K8. The control at Step K5 is to close the variable notch valves, so the solenoid valves 45 and 46 are turned ON at Step K6. On the other hand, the control at Step K7 is to open the variable notch valves, so the solenoid valves 45 and 46 are turned OFF at Step K8.

Now, the regulator control B8 is described with reference to FIG. 50. The line pressure PL, which is supplied from the regulator valve 60, is used for controlling the skew angles of the pump and motor swash plates. To swivel the swash plates, the higher the pressure of the oil passage which is at the higher pressure, the greater the actuation force required. For this reason, the line pressure PL can be set to a high pressure which can satisfy the actuation force necessary for controlling the swash plates even when the oil passage which is at the higher pressure has the maximum pressure. However, if the line pressure were increased in this way, then the loss of the engine's driving force would be great, and the fuel efficiency would be low because the line pressure PL must be maintained at the high pressure. To solve this problem, the regulator control B8 is performed to set a minimum required line pressure PL so as to improve the fuel efficiency.

In this control flow, a target current ICMDr is determined at Step L1 for the actuation of the first solenoid valve 51 in correspondence with the pressure Ph of the oil passage which is at the higher pressure as shown in FIG. 51. With this target current ICMDr, the first linear solenoid valve 51 is actuated at Step L2 to generate the minimum required line pressure PL as shown in FIG. 52. This minimum line pressure is effective in the presence of the pressure Ph of the oil passage which is at the higher pressure, thus improving the fuel efficiency.

The invention being thus described, it will be obvious that the same may be varied in many ways. Such variations are not to be regarded as a departure from the spirit and scope of the invention, and all such modifications as would be obvious to one skilled in the art are intended to be included within the scope of the following claims. 

What is claimed is:
 1. A hydraulic continuously variable transmission for use on a vehicle, comprising: a hydraulic variable pump, which is driven by an engine; said hydraulic variable pump comprising a variable skew angle type swash plate pump; a hydraulic variable motor, which is driven by oil discharged from said hydraulic variable pump, said hydraulic variable motor driving drive wheels; said hydraulic variable motor comprising a variable skew angle type swash plate motor; shift position detecting means, which detects a shift position that is set with a shift lever operated by a driver; brake operation detecting means which detects operation of a vehicle brake by the driver: vehicle stop detecting means, which detects whether the vehicle is in halt or not; pump capacity controlling means, which controls discharge capacity of said hydraulic variable pump; and motor capacity controlling means, which controls discharge capacity of said hydraulic variable motor; wherein: when said vehicle stop detecting means detects that the vehicle is in halt and said shift position detecting means detects that the shift position is in a forward drive position, said pump capacity controlling means controls the discharge capacity of said hydraulic variable pump to a first pump capacity for forward drive; said first pump capacity for forward drive is the capacity that said hydraulic variable pump discharges to supplement an amount of oil that leaks in the transmission when the vehicle is on an uphill road with said engine idling, thus retaining the vehicle in halt on the uphill road without any braking operation; and said motor capacity controlling means controls the discharge capacity of said hydraulic variable motor to a maximum discharge capacity of said hydraulic variable motor; and when said brake operation detecting means detects the operation of the brake, said pump capacity controlling means controls the discharge capacity of said hydraulic variable pump to a second pump capacity which is larger than zero and smaller than said first pump capacity.
 2. A hydraulic continuously variable transmission for use on a vehicle, comprising: a hydraulic variable pump, which is driven by an engine; said hydraulic variable pump comprising a variable skew angle type swash plate pump; a hydraulic variable motor, which is driven by oil discharged from said hydraulic variable pump, said hydraulic variable motor driving drive wheels; said hydraulic variable motor comprising a variable skew angle type swash plate motor; shift position detecting means, which detects a shift position that is set with a shift lever operated by a driver; brake operation detecting means which detects operation of a vehicle brake by the driver; vehicle stop detecting means, which detects whether the vehicle is in halt or not; pump capacity controlling means, which controls discharge capacity of said hydraulic variable pump; and motor capacity controlling means, which controls discharge capacity of said hydraulic variable motor; wherein: when said vehicle stop detecting means detects that the vehicle is in halt and said shift position detecting means detects that the shift position is in a rearward drive position, said pump capacity controlling means controls the discharge capacity of said hydraulic variable pump to a first pump capacity for rearward drive; said first pump capacity for rearward drive is the capacity that said hydraulic variable pump discharges to supplement an amount of oil that leaks in the transmission when the vehicle is on a downhill road with said engine idling, thus retaining the vehicle in halt on the downhill road without any braking operation; and said motor capacity controlling means controls the discharge capacity of said hydraulic variable motor to a maximum discharge capacity of said hydraulic variable motor; when said brake operation detecting means detects the operation of the brake, said pump capacity controlling means controls the discharge capacity of said hydraulic variable pump to a second pump capacity, which is larger than zero and smaller than said first pump capacity.
 3. A hydraulic continuously variable transmission for use on a vehicle between an engine and a drive shaft operatively connected to drive wheels, comprising: a hydraulic variable pump, which is driven by the engine; said hydraulic variable pump comprising a variable skew angle type swash plate pump; a hydraulic variable motor, which is driven by oil discharged from said hydraulic variable pump, said hydraulic variable motor driving the drive shaft for driving the drive wheels; said hydraulic variable motor comprising a variable skew angle type swash plate motor; shift position detecting means, which detects a shift position that is set with a shift lever operated by a driver; brake operation detecting means, which detects operation of a vehicle brake by the driver; vehicle stop detecting means, which detects whether the vehicle is in halt or not; pump capacity controlling means, which controls discharge capacity of said hydraulic variable pump; and motor capacity controlling means, which controls discharge capacity of said hydraulic variable motor; wherein: when said vehicle stop detecting means detects that the vehicle is stopped and said shift position detecting means detects that the shift position is in a drive position, said pump capacity controlling means controls the discharge capacity of said hydraulic variable pump to a first pump capacity for said drive position; said first pump capacity for said drive position is the capacity that said hydraulic variable pump discharges to supplement an amount of oil that leaks in the transmission when the engine is idling and the vehicle is on a road that is inclined upwardly in the direction of said drive position, thus retaining the vehicle stopped on the inclined road without any braking operation; and said motor capacity controlling means controls the discharge capacity of said hydraulic variable motor to a maximum discharge capacity of said hydraulic variable motor; and when said brake operation detecting means detects the operation of the brake, said pump capacity controlling means controls the discharge capacity of said hydraulic variable pump to a second pump capacity, which is larger than zero and smaller than said first pump capacity.
 4. The continuously variable transmission of claim 1, 2, or 3 further comprising: engine speed detecting means for detecting a rotational speed of the engine; wherein, when said engine speed detecting means detects an engine speed lower than a predetermined speed, said pump capacity controlling means lowers both said first pump capacity and said second pump capacity.
 5. The continuously variable transmission of claim 4 wherein said predetermined speed of the engine is an idling speed of the engine when the shift position of the transmission is in park or neutral.
 6. A method of controlling a variable capacity pump and a variable capacity motor of a hydraulic continuously variable transmission connected between an engine and a drive shaft operatively connected to drive wheels of a vehicle while the vehicle is on an inclined road and the engine of the vehicle is idling, comprising the steps of: detecting that the vehicle is stopped and in a shift position for moving the vehicle up the inclined road; controlling a hydraulic discharge capacity of the pump to a first pump capacity to supplement an amount of oil that leaks in the transmission when the vehicle is on the inclined road with the engine idling for retaining the vehicle in a stopped condition without a braking operation; controlling a hydraulic discharge capacity of the motor to a maximum discharge capacity of the variable capacity motor; detecting an operation of a brake by the driver of the vehicle, and controlling the hydraulic discharge capacity of the variable capacity pump to a second pump capacity larger than zero and less than said first pump capacity when operation of the brake by the driver is detected.
 7. The method of claim 6 further comprising the steps of: detecting a rotational speed of the engine; and wherein, when said engine speed detecting means detects an engine speed lower than a predetermined speed, controlling the discharge capacity of said pump for lowering both said first pump capacity and said second pump capacity.
 8. The method of claim 7 wherein said predetermined speed of the engine is an idling speed of the engine when the shift position of the transmission is in park or neutral. 